A generally employed hydraulic circuit for simultaneously driving a plurality of actuators of a construction machine or the like, has all pressure compensating valves set in accordance with the maximum load pressure of the plurality of actuators when a plurality of operating valves are simultaneously operated, thereby making it possible to distribute the oil flow to the plurality of actuators in accordance with the opening area ratio of each operating valve even if the load pressures of the plurality of actuators differ.
In such a hydraulic circuit, when a plurality of actuators are simultaneously operated, the respective pressure compensating valves are set in accordance with the maximum load pressure, and the delivery pressure of the hydraulic pump becomes slightly higher than the set pressure. Therefore, the pressure loss in a pressure compensating valve of an actuator, wherein there is a significant difference between the maximum load pressure and the load pressure of the actuator, becomes high, resulting in a great horsepower loss for the prime mover which drives the hydraulic pump and an increased temperature of the hydraulic oil with consequent quickened deterioration of the hydraulic oil. For instance, when pressurized oil is supplied to a boom cylinder and a swinging motor of a power shovel to lower the boom and to swing the upper body of the power shovel at the same time, the load pressure of the boom is low because the boom lowers automatically due to gravity, while the load pressure of the swinging motor is high because the motor starts and accelerates the upper body, and the pump delivery pressure becomes slightly higher than a set pressure. Therefore, the difference between the high load pressure of the swing motor and the low load pressure of the boom cylinder causes a pressure loss in the pressure compensating valve on the boom cylinder side, resulting in a great pressure loss since the difference in load pressures is great as previously mentioned.
A prior art hydraulic circuit is shown in FIG. 12. More specifically, in FIG. 12, a delivery circuit 1a of a left pump 1 and a delivery circuit 1a of a right pump 1 can be connected together through a flow merging and branching valve 176. A load pressure introduction conduit 178a of a pressure compensating valve 5 of an operating valve 2, which is connected to the delivery circuit 1a of the left pump 1, and a load pressure introduction conduit 178b of the pressure compensating valve 5 of the operating valve 2, which is connected to the delivery circuit 1a of the right pump 1, are connected through a flow merging and branching valve 177. The delivery circuits 1a, 1a and the load introduction conduits 178a, 178b, are made independent by the flow merging and branching valves 176, 177 so that when an actuator 3a, connected to the left delivery circuit 1a, and an actuator 3b, connected to the right delivery circuit 1a, are operated simultaneously, the pressure compensating valve 5 is set in accordance with the load pressures of both actuators, thereby reducing the pressure loss.
In such a hydraulic circuit, however, delivery pressures P.sub.1, P.sub.2 of the pumps are controlled based on a load pressure P.sub.LS, so that they are slightly higher than the load pressure. Therefore, merely providing the delivery passages 1a, 1a of the pumps with relief valves may not allow the hydraulic circuit to exhibit the function thereof. An example of a solution for such problem is a hydraulic circuit, shown in FIG. 13, wherein the delivery circuits 1a, 1a of the right and left pumps 1, 1 are provided with unload valves 170, and the load pressure introduction passages 178a, 178b are provided with relief valves 171 so that the unload valves 170 are unloaded when the pump delivery pressure P.sub.1 exceeds the load pressure P.sub.LS by the aforesaid set pressure or more.
In such a hydraulic circuit, when the load pressure P.sub.LS, of the load pressure introduction passages 178a, 178b exceeds the set pressure of the relief valve 171, the relief valve 171 carries out a relieving operation, causing a part of the load pressure P.sub.LS to flow out into a reservoir tank. When the load pressure P.sub.LS decreases to cause the differential pressure between itself and the pump delivery pressure P.sub.1 to exceed the aforesaid set pressure, the unload valve 170 unloads so as to let a part of the pump delivery pressure P.sub.1 flow into the tank, thus limiting the maximum pump delivery pressure.
However, in the hydraulic circuit shown in FIG. 13, the maximum pump delivery pressure, which is produced when the delivery pressure oils of the right and left pumps 1, 1 are merged and supplied to the actuators, is different from that at the time of flow branching when the delivery pressure oil of either the left pump 1 or the right pump 1 is independently supplied to the actuators.
Specifically, at the time of flow merging, the right and left load pressure introduction passages 178a, 178b are connected, and therefore, when the load pressure P.sub.LS exceeds the set pressure of the relief valve 171, a part of the load pressure from the two relief valves 171, 171 flows out into the tank. At the time of flow branching, the first and second load pressure introduction passages 178a, 178b are separated; and therefore, when the load pressure P.sub.LS, exceeds the set pressure of the relief valves 171, a part of the load pressure flows into the tank through only one relief valve 171. The relief flow rate for the load pressure P.sub.LS will be as shown in FIG. 14, resulting in a different load pressure when the differential pressure (between the pump delivery pressure and the load pressure), at which the unload valve 170 unloads, is reached.
For example, at the time of flow merging, the relief flow rate is large, as shown by a in FIG. 14, since the load pressure is released from both right and left relief valves 171, 171, and therefore a predetermined relief flow rate is reached at a low load pressure P.sub.LS1 . At the time of flow branching, a load pressure P.sub.LS2, at which the aforesaid relief flow rate is reached, becomes high because the relief flow rate is determined by the override characteristic (b of FIG. 14) of the relief valve 171.
Hence, the maximum pump delivery pressure of the pump at the time of flow merging equals P.sub.LS1 plus the unload differential pressure, while the maximum pump delivery pressure at the time of flow branching equals P.sub.LS2 plus the unload differential pressure, the maximum pump delivery pressure at the time of flow branching being higher by P.sub.LS2 -P.sub.LS1. This poses a first problem in that, if the rated pressure is set to the maximum pump delivery pressure at the time of flow merging when designing the hydraulic circuit, then the rated pressure is exceeded at the time of flow branching, adversely affecting the service life of the hydraulic equipment, including pumps, circuits, and actuators; if the rated pressure is set at the maximum pump delivery pressure at the time of flow branching, then the rated pressure is not reached at the time of flow merging, leading to deteriorated performance of the actuators.
Another known hydraulic circuit, which has a pressure compensating valve, for use in a construction machine is disclosed in Japanese Published Unexamined Patent Application (A) 62-88803. More specifically, a hydraulic circuit is known wherein a plurality of closed-center operating valves are disposed in the delivery passage of the pump, pressure compensating valve are provided in a connecting circuit of the operating valves and the actuators, the maximum pressure at the load pressure of each actuator is detected by a shuttle valve or a check valve, and the load pressure is supplied to a spring compartment of each pressure compensating valve, thus producing the set pressure for that particular load pressure.
In such a hydraulic circuit, the pressure compensation is set in accordance with the maximum pressure at the load pressure of the plurality of actuators when the plurality of operating valves are operated simultaneously, thus making it possible to distribute the rate of flow to the plurality of actuators in accordance with the opening area ratios of the operating valves even if the load pressures of the plurality of actuators differ.
Such a hydraulic circuit is provided with a load sensing (LS) valve which controls the tilt of a swash plate of a hydraulic pump in accordance with the differential pressure between the pump delivery pressure of the hydraulic pump and the load pressure, so that the pump delivery pressure is higher than the load pressure only by the aforesaid differential pressure.
On the other hand, if the actuator reaches a stroke end or a significantly high load is applied to the actuator with a resultant extremely high load pressure, then the pump delivery pressure increases accordingly, adversely affecting the service life of the hydraulic equipment. For this reason, the pump delivery pressure is permitted to flow to the reservoir tank, namely, it is relieved, in order to limit the pump delivery pressure.
At the time of the relief, the differential pressure between the pump delivery pressure and the load pressure tends to grow larger than the differential pressure set by the LS valve mentioned above; therefore, the swash plate of the hydraulic pump is automatically set at a minimum swash plate angle, that is, it is set to a cutoff state to minimize the discharge of the hydraulic pump. This causes the relief flow rate to decrease, permitting a reduction in the relief loss.
There is a second problem, however, in that there is normally a time delay of 0.2 to 0.4 second before the swash plate reaches the minimum swash plate angle when the aforesaid load pressure drops from the extremely high level to the low level at which the relief is no longer carried out, and a shortage in the discharge of the hydraulic pump occurs during the delay period.
For example, in the case of the hydraulic circuit for an excavator, such as a power shovel, when a rock is dug up by a bucket during excavation, the load on a bucket cylinder increases markedly and the relieving operation is carried out to set the hydraulic pump at the minimum swash plate angle; when the bucket scoops the rock, which has been dug up, the load on the bucket cylinder grows smaller, causing the relieving operation to stop; and at this time, there is a time delay before the swash plate completes the move from the minimum swash plate angle to a required swash plate angle, resulting in inadequate power to the bucket. This makes an operator erroneously feel that the excavator has a smaller excavating power than its true power.
Further in FIG. 12, a load pressure detecting port 2a of each of the operating valves 2 communicates with a plurality of shuttle valves 160, and the load pressures of the actuators 3a and 3b are compared to detect the higher load pressure.
An example of a known configuration of the plurality of shuttle valves 160 mentioned above is disclosed in the Japanese Published Unexamined Patent Application (A) 1-216174 and is shown in FIG. 15.
Specifically, in FIG. 15, a plurality of mounting holes 181 and communicating holes 182 for communicating adjoining mounting holes 181 are formed in a valve main body 180. Shuttle valve main bodies 183 are fitted into the mounting holes 181. A first inlet port 184, which opens to the mounting hole 181, and a second inlet port 185, which opens to one communicating hole 182, and a second outlet port 186, which opens to the other communicating hole 182, are formed in the shuttle valve main body 183. A ball 187, which communicates either the first inlet port 184 or the second inlet port 185 with the outlet port 186 by means of a working fluid is provided, thus constituting the shuttle valve 160.
In such a shuttle valve 160, when the pressure of the working fluid flowing into the left communicating hole 182 is P.sub.0, the pressure of the working fluid flowing into the first inlet port 184 of the left shuttle valve 160 is P.sub.1, and the pressure of the working fluid flowing into the first inlet port 184 of the right shuttle valve 160 is P.sub.2, then P.sub.0 &lt;P.sub.1 &lt;P.sub.2, the pressure of the working fluid flowing into the first inlet port 184 of the right shuttle valve 160, which is the maximum pressure P.sub.2, is outputted to the right communicating hole ! 82.
However, a third problem is posed by the shuttle valve apparatus mentioned above in that the communicating hole 182 crosses with mounting hole 181 at a right angle, and the communicating hole 182 and the second inlet port 185 are communicated, as shown in FIG. 16, through an axial notch 189 which is formed in the periphery of the shuttle valve main body 183; therefore, the dimension of the shuttle valve main body 183 prevents the communicating area from being made large enough, leading to a large pressure loss. In addition, the working fluid on the first inlet port 184 side and the working fluid of the communicating hole 182 are sealed except for the gap between the mounting hole 181 and the shuttle valve main body 183, and the sealing effect is inadequate, causing the working fluid, which is outputted, to leak toward the lower pressure side. This leakage of the outputted working fluid and the large pressure loss mentioned above adversely affect the control accuracy.
Furthermore, to supply the delivery pressure oil of a single hydraulic pump to a plurality of hydraulic actuators, a delivery passage from the hydraulic pump is provided with a plurality of operating valves and the pressure oil is supplied to all of the hydraulic actuators by switching the operating valves. In this method, however, when the pressurized oil is supplied simultaneously to the plurality of hydraulic actuators, the pressurized oil is supplied only to the hydraulic actuators with smaller load and the pressurized oil is not supplied to the hydraulic actuators with larger load.
A pressure-compensating hydraulic circuit, disclosed in Japanese Published Examined Patent Application (B2) 2-49405, has been proposed as a solution for the problem described above.
The schematic diagram of a pressure-compensating hydraulic circuit is shown in FIG. 17. Specifically, the delivery passage 1a of the hydraulic pump 1 is provided with a plurality of operating valves 2; and the circuits 4, which connect the operating valves 2 and hydraulic actuators 3, are provided with pressure compensating valves 5. The pressure of each circuit 4, that is, the maximum pressure under the load pressure, is detected by a load pressure detecting passage 161 through check valves 160a, 160a, and the detected load pressure is caused to act on a pressure receiving section 5a of each pressure compensating valve 5 in order to set a pressure, which matches the load pressure, to equalize the outlet pressures of all of the operating valves 2 so that, when the operating valves 2 are actuated at the same time, the pressurized oil can be supplied to all the hydraulic actuators 3 according to flow branching ratios, which are proportional to the opening areas of the individual operating valves 2.
In such a pressure-compensating hydraulic circuit, the function of the pressure compensating valves 5 allows the rate of flow to be distributed in proportion to the opening areas of the operating valves 2 independently of the magnitude of the load of the hydraulic actuators 3, making it possible to distribute and supply the delivery pressure oil of the single hydraulic pump 1 to the respective hydraulic actuators 3 in proportion to the operating amounts of the operating valves 2.
According to the hydraulic circuit described above, each pressure compensating valve 5 is set in accordance with the highest load pressure as previously mentioned; therefore, when, for instance, the pressure oil is supplied simultaneously to the power shovel swinging motor and the boom cylinder to lift the boom while swinging the upper body of the power shovel, the starting torque of the swinging motor grows markedly high in the early stage of swinging, causing an extremely high load pressure, and all the pressure compensating valves 5 are set under that extremely high load pressure. This leads to a reduction in the passing flow rate, and the flow which can be supplied to the boom cylinder decreases accordingly, causing a significant decrease in the boom lifting speed and also an increase in the pressure loss in the pressure compensating valves 5.
For this reason, a circuit connected to the load pressure detector of the swinging motor was conventionally provided with a switch valve. The switch valve is closed when supplying the pressurized oil to an actuator other than the swinging motor, so that the load pressure of the swinging motor is not detected, ensuring that the pressure compensating valves are set in accordance with the load pressures of other actuators.
However, there is a fourth problem in that, if the switch valve is opened or closed to detect or not to detect the load pressure of the swinging motor, then the maximum load pressure suddenly changes each time the switch valve is opened or closed. This in turn causes the set pressures of the pressure compensating valves to change abruptly and the passing flow rate to also change suddenly, leading to a sudden change in the flow rate into the actuators with resultant occurrence of a shock.
Further, the applicant proposed in the past the load pressure detector in the pressure-compensating hydraulic circuit shown in FIGS. 18 and 19. As shown in FIGS. 18 and 19, a valve main body 130 is formed with a pair of output ports 132, 132, a pair of pump ports 133, 133, a pair of actuator ports 134,134, and a pair of tank ports 135, 135 located on right and left sides of a spool hole 131a. A load pressure detection port 136 is formed at a lateral midpoint of the spool hole 131a. The load pressure detection port 136 is connected by a check valve 137 to a load pressure detection port 138, which serves as the load pressure detecting circuit. A pair of first small-diameter sections 140, 140 and a pair of second small-diameter sections 141, 141 are formed on right and left sides of a spool 139a, which is fitted in the spool hole 131a, to constitute an operating valve 2. An output port 132 is connected by a pressure compensating valve 5 to the respective actuator port 134. The spool 139a is provided with a pair of right and left load pressure detecting holes 142, 142, with each load pressure detecting hole 142 being opened to a first small-diameter section 140 through a first port 143, opened to a load pressure detection port 136 through a second port 144, and opened to an actuator port 134 through a third port 146 via a ball 145. As shown in FIG. 19, when the spool 19a is in the operating position, the pressure of one output port 132 can flow as the load pressure into a load pressure detection port 136 through a first port 143, a load pressure detecting hole 142, and a second port 144.
However, in a configuration such as shown in FIG. 19, a part of the load pressure, which flows into the load pressure detection port 136 when the operating stroke of the spool 139a is small, flows out into the tank port 135 on the opposite side through a circular groove 147, a groove 147a, a second port 144, the load pressure detecting hole 142, the ball 145, and a drain port 148. Therefore, the detected load pressure is decreased below the pressure of the output port 132, and the delivery flow rate of the hydraulic pump 1 decreases, thus making it possible to prevent the shock in the fine operating area of the operating valve 2. In other words, the delivery flow rate of the hydraulic pump 1 is controlled by utilizing the pump delivery pressure so that the differential pressure between the pump delivery pressure and the load pressure stays constant. As the detected load pressure decreases, the delivery flow rate decreases, and therefore, the delivery flow rate can be reduced by decreasing the load pressure.
However, there is a fifth problem in that, when the spool 139a is located in a neutral position shown in FIG. 18, the pressurized oil of the actuator port 134 based on the holding pressure of the hydraulic actuator 3 leaks not only through the clearance gap between the spool hole 131 and the spool 139a into the pump port 133 and the tank port 135, as shown by the arrows, but also through the third port 146 to the tank port 135 via the drain port 148, as shown by the arrows, causing the pressurized oil on the holding pressure side of the hydraulic actuator 3 to flow out, and the hydraulic actuator 3 may be moved by an external force.
In addition, as the hydraulic circuit for simultaneously driving a plurality of actuators such as in a construction machine or the like, there has been proposed, for example, in Japanese Published Examined Patent Application (B2) 2-49405, a hydraulic circuit as shown in FIG. 17, wherein all pressure compensating valves are set in accordance with the maximum load pressure of the plurality of actuators when a plurality of operating valves are actuated at the same time, and the flow rate is distributed to the plurality of actuators in accordance with the opening area ratios of the operating valves even if the load pressures of the plurality of actuators differ.
In FIG. 17, the delivery passage 1a of the hydraulic pump 1 is provided with a plurality of operating valves 2; and the circuits 4, which connect the operating valves 2 and the hydraulic actuators 3, are provided with the pressure compensating valves 5. The pressure of each circuit 4, that is, the maximum pressure under the load pressure, is detected by a load pressure detecting passage 161 through check valves 160a, and the detected load pressure is caused to act on the pressure receiving section 5a of each pressure compensating valve 5 in order to set a pressure, which matches the load pressure, to equalize the outlet pressures of the operating valves 2 so that, when the operating valves 2 are actuated at the same time, the pressurized oil can be supplied to the hydraulic actuators 3 according to the flow branching ratios, which are proportional to the opening areas of the individual operating valves 2. In such a hydraulic circuit, the function of the pressure compensating valves 5 allows the flow rate to be distributed in proportion to the opening areas of the operating valves 2 independently of the magnitude of the load of the hydraulic actuators 3, making it possible to distribute and supply the delivery pressure oil of the single hydraulic pump 1 to the respective hydraulic actuators 3 in proportion to the operating amounts of the operating valves 2.
The pressure compensating valves 5 of FIG. 17 are, however, pressed toward the closing side by the maximum load pressure P.sub.LS, acting on the first pressure receiving section 5a, and the spring force, while it is pressed toward the opening side by an operating valve outlet pressure (flow pressure after meter in) P.sub.0 acting on the second pressure receiving section 5b; the pressure receiving areas of the first pressure receiving section 5a and the second pressure receiving section 5b are the same; hence the operating valve outlet pressure P.sub.0 equals the maximum load pressure P.sub.LS. Thus, if the hydraulic actuators are the right and left travelling hydraulic motors, the opening of the left operating valve 2 is set large for the driving side and the opening of the right operating valve 2 is set small for the driven side for swinging and traveling, and the load pressure P.sub.LS1 acting on the first pressure receiving section 5a of the pressure compensating valve 5 on the braking side is equal to the load pressure P.sub.LS2 on the driving side, then the flow rate supplied to the right and left traveling hydraulic motors will be the values which are proportional to the opening ratios of both operating valves 2, and the swing radius is controlled by the opening ratios of the two operating valves 2.
However, there is a sixth problem in that the load pressure P.sub.LS acting on the first pressure receiving section 5a of the pressure compensating valve 5 on the braking side incurs a pressure loss due to leakage or the like, and it becomes lower than the load pressure P.sub.LS on the driving side; this causes the inflow into the traveling hydraulic motor on the braking side to increase and the turning radius to be different from the value, which matches the opening ratios of both operating valves 2, making the turning and traveling difficult.
There is also a seventh problem when the hydraulic circuit of FIG. 20 is used for a power shovel, the two left hydraulic actuators 3, 3 act as the traveling motors while the right single hydraulic actuator 3 acts as the boom cylinder, and the two left operating valves 2, 2 are operated to supply the pressurized oil to the traveling motors for traveling. If the right operating valve 2 is operated to supply the pressurized oil to the right boom cylinder to raise the boom, then the rate of flow into the right and left traveling motors suddenly decreases and a traveling deceleration shock occurs, since the rates of flow into the right and left traveling motors 3, 3 and to the boom cylinder 3 are divided in accordance with the opening area ratios of the respective operating valves 2.
If, for instance, a total opening area of the two operating valves 2 for traveling is equal to the opening area of the operating valve 2 for the boom, then the traveling speed will be reduced to half that of independent operation and the boom lifting speed will be also reduced to half that of independent operation.